JP4776447B2 - Variable valve operating device for internal combustion engine - Google Patents

Variable valve operating device for internal combustion engine Download PDF

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Publication number
JP4776447B2
JP4776447B2 JP2006161760A JP2006161760A JP4776447B2 JP 4776447 B2 JP4776447 B2 JP 4776447B2 JP 2006161760 A JP2006161760 A JP 2006161760A JP 2006161760 A JP2006161760 A JP 2006161760A JP 4776447 B2 JP4776447 B2 JP 4776447B2
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Prior art keywords
variable
engine
valve
lift
cranking
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Expired - Fee Related
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JP2006161760A
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Japanese (ja)
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JP2007332780A (en
Inventor
信 中村
誠之助 原
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日立オートモティブシステムズ株式会社
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    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01LCYCLICALLY OPERATING VALVES FOR MACHINES OR ENGINES
    • F01L1/00Valve-gear or valve arrangements, e.g. lift-valve gear
    • F01L1/34Valve-gear or valve arrangements, e.g. lift-valve gear characterised by the provision of means for changing the timing of the valves without changing the duration of opening and without affecting the magnitude of the valve lift
    • F01L1/344Valve-gear or valve arrangements, e.g. lift-valve gear characterised by the provision of means for changing the timing of the valves without changing the duration of opening and without affecting the magnitude of the valve lift changing the angular relationship between crankshaft and camshaft, e.g. using helicoidal gear
    • F01L1/3442Valve-gear or valve arrangements, e.g. lift-valve gear characterised by the provision of means for changing the timing of the valves without changing the duration of opening and without affecting the magnitude of the valve lift changing the angular relationship between crankshaft and camshaft, e.g. using helicoidal gear using hydraulic chambers with variable volume to transmit the rotating force
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01LCYCLICALLY OPERATING VALVES FOR MACHINES OR ENGINES
    • F01L1/00Valve-gear or valve arrangements, e.g. lift-valve gear
    • F01L1/02Valve drive
    • F01L1/022Chain drive
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01LCYCLICALLY OPERATING VALVES FOR MACHINES OR ENGINES
    • F01L13/00Modifications of valve-gear to facilitate reversing, braking, starting, changing compression ratio, or other specific operations
    • F01L13/0015Modifications of valve-gear to facilitate reversing, braking, starting, changing compression ratio, or other specific operations for optimising engine performances by modifying valve lift according to various working parameters, e.g. rotational speed, load, torque
    • F01L13/0021Modifications of valve-gear to facilitate reversing, braking, starting, changing compression ratio, or other specific operations for optimising engine performances by modifying valve lift according to various working parameters, e.g. rotational speed, load, torque by modification of rocker arm ratio
    • F01L13/0026Modifications of valve-gear to facilitate reversing, braking, starting, changing compression ratio, or other specific operations for optimising engine performances by modifying valve lift according to various working parameters, e.g. rotational speed, load, torque by modification of rocker arm ratio by means of an eccentric
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02DCONTROLLING COMBUSTION ENGINES
    • F02D13/00Controlling the engine output power by varying inlet or exhaust valve operating characteristics, e.g. timing
    • F02D13/02Controlling the engine output power by varying inlet or exhaust valve operating characteristics, e.g. timing during engine operation
    • F02D13/0223Variable control of the intake valves only
    • F02D13/0226Variable control of the intake valves only changing valve lift or valve lift and timing
    • F02D13/023Variable control of the intake valves only changing valve lift or valve lift and timing the change of valve timing is caused by the change in valve lift, i.e. both valve lift and timing are functionally related
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01LCYCLICALLY OPERATING VALVES FOR MACHINES OR ENGINES
    • F01L1/00Valve-gear or valve arrangements, e.g. lift-valve gear
    • F01L1/02Valve drive
    • F01L1/04Valve drive by means of cams, camshafts, cam discs, eccentrics or the like
    • F01L1/047Camshafts
    • F01L1/053Camshafts overhead type
    • F01L2001/0537Double overhead camshafts [DOHC]
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01LCYCLICALLY OPERATING VALVES FOR MACHINES OR ENGINES
    • F01L1/00Valve-gear or valve arrangements, e.g. lift-valve gear
    • F01L1/34Valve-gear or valve arrangements, e.g. lift-valve gear characterised by the provision of means for changing the timing of the valves without changing the duration of opening and without affecting the magnitude of the valve lift
    • F01L1/344Valve-gear or valve arrangements, e.g. lift-valve gear characterised by the provision of means for changing the timing of the valves without changing the duration of opening and without affecting the magnitude of the valve lift changing the angular relationship between crankshaft and camshaft, e.g. using helicoidal gear
    • F01L1/3442Valve-gear or valve arrangements, e.g. lift-valve gear characterised by the provision of means for changing the timing of the valves without changing the duration of opening and without affecting the magnitude of the valve lift changing the angular relationship between crankshaft and camshaft, e.g. using helicoidal gear using hydraulic chambers with variable volume to transmit the rotating force
    • F01L2001/34423Details relating to the hydraulic feeding circuit
    • F01L2001/34426Oil control valves
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01LCYCLICALLY OPERATING VALVES FOR MACHINES OR ENGINES
    • F01L1/00Valve-gear or valve arrangements, e.g. lift-valve gear
    • F01L1/34Valve-gear or valve arrangements, e.g. lift-valve gear characterised by the provision of means for changing the timing of the valves without changing the duration of opening and without affecting the magnitude of the valve lift
    • F01L1/344Valve-gear or valve arrangements, e.g. lift-valve gear characterised by the provision of means for changing the timing of the valves without changing the duration of opening and without affecting the magnitude of the valve lift changing the angular relationship between crankshaft and camshaft, e.g. using helicoidal gear
    • F01L1/3442Valve-gear or valve arrangements, e.g. lift-valve gear characterised by the provision of means for changing the timing of the valves without changing the duration of opening and without affecting the magnitude of the valve lift changing the angular relationship between crankshaft and camshaft, e.g. using helicoidal gear using hydraulic chambers with variable volume to transmit the rotating force
    • F01L2001/3445Details relating to the hydraulic means for changing the angular relationship
    • F01L2001/34453Locking means between driving and driven members
    • F01L2001/34469Lock movement parallel to camshaft axis
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01LCYCLICALLY OPERATING VALVES FOR MACHINES OR ENGINES
    • F01L1/00Valve-gear or valve arrangements, e.g. lift-valve gear
    • F01L1/34Valve-gear or valve arrangements, e.g. lift-valve gear characterised by the provision of means for changing the timing of the valves without changing the duration of opening and without affecting the magnitude of the valve lift
    • F01L1/344Valve-gear or valve arrangements, e.g. lift-valve gear characterised by the provision of means for changing the timing of the valves without changing the duration of opening and without affecting the magnitude of the valve lift changing the angular relationship between crankshaft and camshaft, e.g. using helicoidal gear
    • F01L1/3442Valve-gear or valve arrangements, e.g. lift-valve gear characterised by the provision of means for changing the timing of the valves without changing the duration of opening and without affecting the magnitude of the valve lift changing the angular relationship between crankshaft and camshaft, e.g. using helicoidal gear using hydraulic chambers with variable volume to transmit the rotating force
    • F01L2001/3445Details relating to the hydraulic means for changing the angular relationship
    • F01L2001/34483Phaser return springs
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01LCYCLICALLY OPERATING VALVES FOR MACHINES OR ENGINES
    • F01L13/00Modifications of valve-gear to facilitate reversing, braking, starting, changing compression ratio, or other specific operations
    • F01L13/0015Modifications of valve-gear to facilitate reversing, braking, starting, changing compression ratio, or other specific operations for optimising engine performances by modifying valve lift according to various working parameters, e.g. rotational speed, load, torque
    • F01L13/0063Modifications of valve-gear to facilitate reversing, braking, starting, changing compression ratio, or other specific operations for optimising engine performances by modifying valve lift according to various working parameters, e.g. rotational speed, load, torque by modification of cam contact point by displacing an intermediate lever or wedge-shaped intermediate element, e.g. Tourtelot
    • F01L2013/0073Modifications of valve-gear to facilitate reversing, braking, starting, changing compression ratio, or other specific operations for optimising engine performances by modifying valve lift according to various working parameters, e.g. rotational speed, load, torque by modification of cam contact point by displacing an intermediate lever or wedge-shaped intermediate element, e.g. Tourtelot with an oscillating cam acting on the valve of the "Delphi" type
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01LCYCLICALLY OPERATING VALVES FOR MACHINES OR ENGINES
    • F01L2800/00Methods of operation using a variable valve timing mechanism
    • F01L2800/01Starting
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02DCONTROLLING COMBUSTION ENGINES
    • F02D13/00Controlling the engine output power by varying inlet or exhaust valve operating characteristics, e.g. timing
    • F02D13/02Controlling the engine output power by varying inlet or exhaust valve operating characteristics, e.g. timing during engine operation
    • F02D13/0223Variable control of the intake valves only
    • F02D13/0234Variable control of the intake valves only changing the valve timing only
    • F02D13/0238Variable control of the intake valves only changing the valve timing only by shifting the phase, i.e. the opening periods of the valves are constant
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02DCONTROLLING COMBUSTION ENGINES
    • F02D13/00Controlling the engine output power by varying inlet or exhaust valve operating characteristics, e.g. timing
    • F02D13/02Controlling the engine output power by varying inlet or exhaust valve operating characteristics, e.g. timing during engine operation
    • F02D13/0269Controlling the valves to perform a Miller-Atkinson cycle
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02DCONTROLLING COMBUSTION ENGINES
    • F02D13/00Controlling the engine output power by varying inlet or exhaust valve operating characteristics, e.g. timing
    • F02D13/02Controlling the engine output power by varying inlet or exhaust valve operating characteristics, e.g. timing during engine operation
    • F02D2013/0292Controlling the engine output power by varying inlet or exhaust valve operating characteristics, e.g. timing during engine operation in the start-up phase, e.g. for warming-up cold engine or catalyst
    • Y02T10/142
    • Y02T10/18

Description

  The present invention relates to a variable valve operating apparatus for an internal combustion engine that can improve engine starting performance.

  As is well known, the opening / closing timing (valve timing) of an engine valve, particularly an intake valve, at the initial stage of the engine is controlled with high accuracy in accordance with the tightening of exhaust emission regulations at the start of the internal combustion engine and the increase of the restart frequency by the hybrid vehicle. It is hoped that.

  In order to realize these, the variable valve that can control the opening and closing timing of the engine valve variably according to the engine operating condition such as at the start by changing the relative rotation phase of the vane fixed to the camshaft and the timing sprocket by hydraulic pressure Various devices are provided.

  However, such a hydraulic variable valve device has a low control hydraulic pressure at the time of extremely low engine rotation, such as at the time of start-up, and lacks switching power, and at the time of cranking, the reverse of the cranking rotation direction. In addition to being easily held on the corner side, quick switching response to the advance side is deteriorated due to friction of each component.

  Therefore, as one of these countermeasures, for example, a variable valve gear described in Patent Document 1 below is provided.

In this variable valve operating system, a phase conversion mechanism is provided on the intake valve side, and when the engine is stopped, the vane is fixed by a lock pin at a position that has been advanced to some extent, and is started with a certain degree of advance at the time of restart. Is. This is expected to improve start-up performance such as reducing the cold engine emissions by stabilizing the combustion by bringing the valve overlap between the exhaust valve and intake valve close and the closing timing of the intake valve close to bottom dead center. Has been.
JP 2005-233049 A

  However, in the variable valve operating apparatus for an internal combustion engine described in Patent Document 1, the lock pin and the engagement hole in which the lock pin engages under various stop modes of the engine such as a sudden brake or an engine stall of the vehicle. It is difficult to make sure that they are engaged and locked. Therefore, the behavior of the vane after the engine is stopped becomes unstable, and there is a possibility that a reliable emission reduction effect at the start of the cold machine cannot be obtained as in the prior art.

Even if the lock pin is engaged and locked in the engagement hole, the next start-up state is not always cold, and if it is warm, the valve overlap Due to the large size, there is a problem that idling rotation becomes uneasy.

  The present invention has been devised in view of the technical problem of the conventional variable valve operating device, and even when the phase of the second variable mechanism is not suitable for starting when the engine is started, the first variable mechanism is An object of the present invention is to provide a variable valve operating apparatus for an internal combustion engine that can improve the starting performance by forcibly changing the opening / closing timing of the engine valve according to the state.

According to the first aspect of the present invention, a first variable mechanism that can control the operating characteristics of the engine valve during engine stop or cranking, and a second variable that controls the operating state of the engine valve separately from the first variable mechanism. comprising a variable mechanism, and in accordance with the operating state of the cranking start before the second variable mechanism of the engine, cranking before or during cranking, the so as to satisfy a predetermined start condition of the engine A control signal for controlling the operation of the first variable mechanism is output prior to cranking, and cranking is started with a predetermined delay time from the peak value of the control current of the first variable mechanism .

  According to this invention, for example, due to malfunction of the second variable mechanism, the opening / closing timing of the engine valve controlled when the engine is stopped deviates from the target opening / closing timing suitable for starting the engine before cranking. In this case, the opening / closing timing is controlled by changing the characteristic of the engine valve by the first variable mechanism by an electric actuator or the like. As a result, it is possible to obtain a desired starting performance that facilitates cranking rotation, for example.

  The startability of the engine is not limited to the ease of cranking rotation, but the opening / closing timing of the engine valve that reduces emissions of HC (hydrocarbon) during engine combustion may be set as a target set value. Is possible.

  Embodiments of a variable valve operating apparatus for an internal combustion engine according to the present invention will be described below in detail with reference to the drawings. This embodiment shows a so-called four-cycle multi-cylinder internal combustion engine applied to the intake valve side.

  First, the overall configuration of the internal combustion engine according to the present invention will be described with reference to FIG. 1. A piston 01 provided in a cylinder bore formed in a cylinder block SB and slidable up and down, and an interior of the cylinder head SH. And a pair of intake valves 4 per cylinder that are slidably provided on the cylinder head SH and open and close the open ends of the intake and exhaust ports IP and EP, respectively. 4 and exhaust valves 5 and 5.

  The piston 01 is connected to the crankshaft 02 via a connecting rod 03, and forms a combustion chamber 04 between the crown surface and the lower surface of the cylinder head SH.

  A throttle valve SV for controlling the intake air amount is provided in the upstream side of the intake manifold Ia of the intake pipe I connected to the intake port IP, and a fuel injection valve (not shown) is provided on the downstream side. It has been. In addition, a spark plug 05 is provided substantially at the center of the cylinder head SH.

  The crankshaft 02 can be rotated forward and backward by an electric starter motor 07 via a pinion gear mechanism 06.

  The variable means as the variable valve mechanism is, as shown in FIGS. 1 and 2, a variable lift that is a first variable mechanism that controls the valve lift and the operating angle (open period) of both intake valves 4 and 4. A mechanism (VEL) 1 and a valve timing variable mechanism (VTC) 2 which is a second variable mechanism for controlling the lift phase of the intake valves 4 and 4 are configured.

The lift variable mechanism 1 has the same configuration as that described in, for example, Japanese Patent Application Laid-Open No. 2003-172112 filed earlier by the present applicant. A hollow drive shaft 6 that is rotatably supported, a drive cam 7 that is an eccentric rotary cam fixed to the outer peripheral surface of the drive shaft 6 by press-fitting or the like, and an outer peripheral surface of the drive shaft 6 that is swingable. Two swing cams 9, 9 that are supported and slidably contact the upper surfaces of the valve lifters 8, 8 disposed at the upper ends of the intake valves 4, 4 to open the intake valves 4, 4, and drive And a transmission mechanism that is interposed between the cam 7 and the swing cams 9 and 9 and converts the rotational force of the drive cam 7 into swing motion and transmits the swing cams 9 and 9 as swing force. Yes.

  The drive shaft 6 is transmitted with rotational force from the crankshaft 02 by a timing chain (not shown) via a timing sprocket 30 provided at one end, and this rotational direction is clockwise (arrow direction) in FIG. ) Is set.

  The drive cam 7 has a substantially ring shape, and is fixed to the drive shaft 6 through a drive shaft insertion hole formed in the inner axial direction. The shaft center of the cam body extends from the shaft center of the drive shaft 6. Offset by a predetermined amount in the radial direction.

  As shown in FIG. 2 and FIG. 3 and the like, both the swing cams 9 have substantially the same raindrop shape and are integrally provided at both ends of the annular cam shaft 10. A shaft 10 is rotatably supported on the drive shaft 6 via an inner peripheral surface. A cam surface 9a is formed on the lower surface, a base circle surface on the shaft side of the camshaft 10, a ramp surface extending in an arc shape from the base circle surface to the cam nose portion side, and from the ramp surface to the distal end side of the cam nose portion. A lift surface connected to the top surface of the maximum lift is formed, and the base circle surface, the ramp surface, and the lift surface are in contact with predetermined positions on the upper surface of each valve lifter 8 according to the swing position of the swing cam 9. It comes to touch.

  The transmission mechanism includes a rocker arm 11 disposed above the drive shaft 6, a link arm 12 linking the one end 11 a of the rocker arm 11 and the drive cam 7, the other end 11 b of the rocker arm 11, and a swing cam 9. And a link rod 13 that cooperates with each other.

  The rocker arm 11 has a cylindrical base portion at the center thereof rotatably supported by a control cam, which will be described later, via a support hole, and one end portion 11 a is rotatably connected to the link arm 12 by a pin 14. On the other hand, the other end 11 b is rotatably connected to one end 13 a of the link rod 13 via a pin 15.

  The link arm 12 has a fitting hole in which the cam body of the drive cam 7 is rotatably fitted at the center position of a relatively large-diameter annular base 12a, while the protruding end 12b is the pin. 14 is connected to one end 11a of the rocker arm.

  The other end portion 13 b of the link rod 13 is rotatably connected to the cam nose portion of the swing cam 9 via a pin 16.

  A control shaft 17 is rotatably supported by the same bearing member at a position above the drive shaft 6, and is slidably fitted into the support hole of the rocker arm 11 on the outer periphery of the control shaft 17. A control cam 18 serving as a swing fulcrum is fixed.

  The control shaft 17 is disposed in the longitudinal direction of the engine in parallel with the drive shaft 6 and is rotationally controlled by a drive mechanism 19. On the other hand, the control cam 18 has a cylindrical shape, and the axial center position is deviated from the axial center of the control shaft 17 by a predetermined amount.

  The drive mechanism 19 includes an electric motor 20 that is fixed to one end of a housing (not shown), and a ball screw transmission means 21 that is provided inside the housing and transmits the rotational driving force of the electric motor 20 to the control shaft 17. It is composed of

  The electric motor 20 is constituted by a proportional DC motor and is driven by a control signal from a controller 22 which is a control mechanism for detecting an engine operating state.

  The controller 22 outputs an output signal from a crank angle sensor 27 that detects a current engine speed N (rpm) from a crank angle, an intake air amount (load) from an air flow meter, an accelerator opening sensor, a vehicle speed. The current engine operating state is detected from various information signals from a sensor, a gear position sensor, an engine coolant temperature sensor that detects the temperature T1 of the engine body, and the like. Further, a detection signal from the drive shaft angle sensor 28 for detecting the rotation angle of the drive shaft 6 and the atmospheric humidity H1 from the atmospheric humidity sensor are input.

  The ball screw transmission means 21 includes a ball screw shaft 23 disposed substantially coaxially with the drive shaft of the electric motor 20, a ball nut 24 that is a moving member screwed onto the outer periphery of the ball screw shaft 23, and the control It is mainly comprised from the linkage arm 25 connected with the one end part of the axis | shaft 17 along the diameter direction, and the link member 26 which links this linkage arm 25 and the said ball nut 24. As shown in FIG.

  In the ball screw shaft 23, a ball circulation groove having a predetermined width is continuously formed in a spiral shape on the entire outer peripheral surface excluding both end portions, and one end portion is coupled to the drive shaft of the electric motor 20, and by this coupling, The rotational driving force of the electric motor 20 is transmitted to the ball screw shaft 23, and the ball screw shaft 23 is allowed to move slightly in the axial direction.

  The ball nut 24 is formed in a substantially cylindrical shape, and a guide groove for continuously holding a plurality of balls is formed in a spiral manner in cooperation with the ball circulation groove on the inner peripheral surface. An axial moving force is applied to the ball nut 24 while converting the rotational motion of the ball screw shaft 23 into a linear motion via each ball. The ball nut 24 is urged toward the electric motor 20 by the spring force of the coil spring 31 so that the backlash gap between the ball nut 24 and the ball screw shaft 23 disappears.

  Hereinafter, the basic operation of the variable lift mechanism 1 will be described. In a predetermined operation range, the controller 22 is rotationally driven by energization control from the controller 22 to the electric motor 20, and the ball screw shaft 23 is rotated by the rotational torque of the electric motor 20. When rotating in one direction, the ball nut 24 moves linearly in a maximum direction (direction approaching the electric motor 20), whereby the control shaft 17 rotates in one direction via the link member 39 and the linkage arm 25. .

  Accordingly, as shown in FIGS. 3A and 3B (rear view), the control cam 18 rotates about the axis of the control shaft 17 with the same radius, and the thick portion is separated upward from the drive shaft 6. Moving. As a result, the other end portion 11b of the rocker arm 11 and the pivot point of the link rod 13 move upward with respect to the drive shaft 6. Therefore, each swing cam 9 is connected to the cam nose portion side via the link rod 13. The whole is forcibly pulled up and rotated in the counterclockwise direction shown in FIG.

  Therefore, when the drive cam 7 rotates and pushes up the one end portion 11a of the rocker arm 11 via the link arm 12, the lift amount is transmitted to the swing cam 9 and the valve lifter 16 via the link rod 13, thereby As shown by the valve lift curve of FIG. 5, the intake valves 4 and 4 have a small lift (L1), and the operating angle D1 (half of the crank valve opening period) becomes small.

  Therefore, a decompression effect, a small lift low friction effect, a fuel consumption effect, and the like can be obtained.

  In another operating state, when the electric motor 20 is rotated in reverse by a control signal from the controller 22 and this rotational torque is transmitted to the ball screw shaft 23 and rotated, the ball nut 24 is moved in the opposite direction along with this rotation. Move straight to. Thereby, the control shaft 17 is rotationally driven by a predetermined amount in the counterclockwise direction (the direction away from the electric motor 20) in FIG.

  For this reason, the shaft center of the control cam 18 is held at a rotational angle position that is lower than the shaft center of the control shaft 17 by a predetermined amount, and the thick portion moves downward. For this reason, the entire rocker arm 11 moves clockwise from the position of FIG. 3, whereby each swing cam 9 is forcibly pushed down the cam nose portion side via the link member 13, and the entire rocker arm 11 is clockwise. It turns slightly.

  Therefore, when the drive cam 7 rotates and pushes up the one end portion 11a of the rocker arm 11 via the link arm 12, the lift amount is transmitted to each swing cam 9 and the valve lifter 8 via the link member 13, and the intake valve 4 , 4 is a middle lift (L2) as shown in FIG. 5, and the operating angle D2 is also increased. As a result, the closing timing of the intake valves 4 and 4 is controlled in the vicinity of the bottom dead center on the retarded side, so that the effective compression ratio is increased and combustion at the time of cold start is improved. Moreover, the charging efficiency of fresh air is increased and the combustion torque is increased.

  Further, in the low rotation and low load range after the engine has warmed up, for example, if the lift is controlled to the small lift L1 and the retard angle control is performed by the variable valve timing mechanism 2, the valve overlap with the exhaust valves 5 and 5 is small. Thus, combustion is stable, and since the valve friction is small with a small lift, fuel consumption can be improved.

  Further, in the middle load range or the like, if the controller 22 controls the vicinity of the middle lift (L2) as described above and the lift phase is advanced by the valve timing variable mechanism 2, the valve overlap with the exhaust valves 5 and 5 is prevented. It becomes larger and the pumping loss is reduced, so that fuel efficiency is improved.

  Further, in the case of shifting to the high rotation / high load region, the electric motor 20 is further rotated in the reverse direction by the control signal from the controller 22, and the control shaft 17 further rotates the control cam 18 in the counterclockwise direction. , B as shown in FIG. For this reason, the entire rocker arm 11 further moves toward the drive shaft 6, and the other end 11 b presses the cam nose portion of the swing cam 9 downward via the link rod 13, thereby moving the entire swing cam 9. It is further rotated clockwise by a predetermined amount.

  Therefore, when the drive cam 7 rotates and pushes up the one end portion 11a of the rocker arm 11 via the link arm 12, the lift amount is transmitted to the swing cam 9 and the valve lifter 8 via the link rod 13. The valve lift amount continuously increases from L2 to L3 as shown in FIG. As a result, it is possible to increase the intake charging efficiency in the high rotation range, thereby improving the output.

  That is, the lift amount of the intake valves 4 and 4 changes continuously from the small lift L1 to the large lift L3 according to the operating state of the engine. Also changes continuously from a small lift D1 to a large lift D3.

  The valve timing variable mechanism 2 is a so-called vane type, and as shown in FIGS. 6 and 7, a timing sprocket 30 for transmitting a rotational force to the drive shaft 6 and an end portion of the drive shaft 6. A vane member 32 that is fixed and rotatably accommodated in the timing sprocket 30 and a hydraulic circuit 33 that rotates the vane member 32 forward and backward by hydraulic pressure are provided.

  The timing sprocket 30 includes a housing 34 that rotatably accommodates the vane member 32, a disk-shaped front cover 35 that closes the front end opening of the housing 34, and a substantially disk that closes the rear end opening of the housing 34. The housing 34, the front cover 35, and the rear cover 36 are integrally fastened together by four small-diameter bolts 37 from the axial direction of the drive shaft 6.

  The housing 34 has a cylindrical shape in which both front and rear ends are formed, and shoes 34a that are four partition walls project inward at a position of about 90 ° in the circumferential direction of the inner peripheral surface.

  Each of the shoes 34a has a substantially trapezoidal cross section, and four bolt insertion holes 34b through which the shaft portions of the respective bolts 37 are inserted are formed at substantially central positions so as to penetrate in the axial direction. A U-shaped seal member 38 and a leaf spring (not shown) that presses the seal member 38 inwardly are fitted and held in a holding groove that is cut out along the axial direction at the position.

  The front cover 35 is formed in the shape of a disk plate, and a support hole 35a having a relatively large diameter is formed in the center. The front cover 35 is not shown at a position corresponding to each bolt insertion hole of the housing 34 on the outer periphery. These four bolt holes are drilled.

  The rear cover 36 is integrally provided with a gear portion 36a meshing with the timing chain on the rear end side, and a large-diameter bearing hole 36b is formed in the axial direction so as to penetrate therethrough.

  The vane member 32 includes an annular vane rotor 32a having a bolt insertion hole in the center, and four vanes 32b that are integrally provided at approximately 90 ° in the circumferential direction of the outer peripheral surface of the vane rotor 32a.

  In the vane rotor 32a, a small-diameter cylindrical portion on the front end side is rotatably supported by the support hole 35a of the front cover 35, while a small-diameter cylindrical portion on the rear end side is freely rotatable in the bearing hole 36b of the rear cover 36. It is supported.

  The vane member 32 is fixed to the front end portion of the drive shaft 6 from the axial direction by a fixing bolt 39 inserted through the bolt insertion hole of the vane rotor 32a from the axial direction.

  Three of the vanes 32b are formed in a relatively long and narrow rectangular shape, and the other one is formed in a relatively large trapezoidal shape. The three vanes 32b are substantially the same in width and length. In contrast, the width of one vane 32b is set to be larger than that of the three vanes 32, and the weight balance of the entire vane member 32 is achieved.

  Each vane 32b is disposed between the shoes 34a and has a U-shaped seal member 40 slidably contacting the inner peripheral surface of the housing 34 in an elongated holding groove formed in the axial direction of each outer surface. Leaf springs that press the seal member 40 toward the inner peripheral surface of the housing 34 are fitted and held, respectively. Further, two substantially circular concave grooves 32c are formed on one side surface of each vane 32b opposite to the rotation direction of the drive shaft 6, respectively.

  Further, four advance chambers 41, which are advance chambers, and retard chambers 42, which are retard chambers, are respectively formed between both sides of each vane 32b and both sides of each shoe 34a.

  As shown in FIG. 6, the hydraulic circuit 33 includes a first hydraulic passage 43 that supplies and discharges hydraulic oil pressure to and from each advance chamber 41, and hydraulic oil pressure to each retard chamber 42. The two hydraulic passages 43 and 44 are provided with a supply passage 45 and a drain passage 46 via an electromagnetic switching valve 47 for switching the passage. Connected. The supply passage 45 is provided with a one-way oil pump 49 for pumping oil in the oil pan 48, while the downstream end of the drain passage 46 communicates with the oil pan 48.

  The first and second hydraulic passages 43 and 44 are formed in a cylindrical passage constituting portion 39, and one end of the passage constituting portion 39 extends from the small diameter cylindrical portion of the vane rotor 32a to the inside of the support hole 32d. The other end portion is connected to the electromagnetic switching valve 47.

  Further, between the outer peripheral surface of one end portion of the passage constituting portion 39 and the inner peripheral surface of the support hole 14d, three annular seal members 27 for separating and sealing one end side of each of the hydraulic passages 43 and 44 are fitted. It is fixed.

  The first hydraulic passage 43 is formed in an oil chamber 43a formed at an end of the support hole 32d on the drive shaft 6 side, and is formed almost radially inside the vane rotor 32a. And four branch paths 43b communicating with each other.

  On the other hand, the second hydraulic passage 44 is stopped in one end portion of the passage constituting portion 39, and is bent into a substantially L shape inside the annular chamber 44a formed on the outer peripheral surface of the one end portion and the vane rotor 32. The annular chamber 44a and a second oil passage 44b communicating with each retarded angle chamber 42 are provided.

  The electromagnetic switching valve 47 is a four-port three-position type, and an internal valve body is configured to relatively switch and control each of the hydraulic passages 43 and 44, the supply passage 45, and the drain passage 46, Switching operation is performed by a control signal from the controller 22.

  This controller 22 is common to the variable lift mechanism 1 and detects the engine operating state, and the relative rotation between the timing sprocket 30 and the drive shaft 6 by signals from the crank angle sensor 27 and the drive shaft angle sensor 28. The position is detected.

  By switching to the neutral position of the electromagnetic switching valve 47, hydraulic oil is not actively supplied to the advance chamber 41 and the retard chamber 42 when the engine is started.

  Further, a pair of coil springs that are biasing means for biasing the vane member 32 to advance toward the advance side are provided between one side surface of each vane 32b and the facing surface 10b of each shoe 34a facing the one side surface. 55 and 56 are arranged, respectively.

  The two coil springs 55 and 56 appear to overlap each other in FIGS. 7 and 8, but in actuality, they are formed independently of each other and formed in parallel with each other, and each axial length The length (coil length) is set larger than the length between one side surface of the vane 32b and the opposing surface of the shoe 34a, and both are set to the same length.

  The coil springs 55 and 56 are arranged in parallel with an inter-axis distance that does not contact each other at the time of maximum compression deformation, and a thin plate-like retainer (not shown) whose one end fits into the groove 32c of each shoe 34a. Are connected through. Further, the spring force of both the coil springs 55 and 56 is set to be relatively small, and at the time of cranking, the vane member 32 is left behind in the retarded direction due to the sliding resistance of the vane member 32. It is set so that it can be pushed back to the middle position.

  Hereinafter, the basic operation of the variable valve timing mechanism 2 will be described.

  First, when the engine is stopped, the output of the control current from the controller 22 to the electromagnetic switching valve 47 is stopped, and the valve body communicates the supply passage 45 and the second hydraulic passage 44 on the retard side. However, when the engine speed becomes zero, the oil pressure of the oil pump 49 does not act and the supply oil pressure becomes zero.

  Here, since the timing sprocket 30 rotates in the clockwise direction in FIG. 7 immediately before the engine stops, the vane member 32 will be left behind (counterclockwise) by the friction with the housing 34. And Here, the spring force of each of the coil springs 55 and 56 is biased toward the advance side, but this biasing force is relatively weak and does not cause the vane member 32 to rotate to the most retarded angle. The vane member 32 is stably held near the intermediate position from the retard side.

Next, when the engine is started, that is, when the ignition switch is turned on to rotate the starter motor 07 to crank the crankshaft 02, the vane member 32 is held near the intermediate position at this time. Therefore, the closing timing (IVC) approaches the bottom dead center, the effective compression ratio becomes high, and good startability can be secured. Also, because there is a moderate valve overlap, cold machine emissions are improved.

  Thereafter, when the warm-up proceeds, the electromagnetic switching valve 47 causes the supply passage 45 and the second hydraulic passage 44 to communicate with each other and the drain passage 46 and the first hydraulic passage 43 communicate with each other according to the control signal output from the controller 22. For this reason, the hydraulic pressure pumped from the oil pump 49 is supplied to the retard chamber 42 through the second hydraulic passage 44, while the hydraulic chamber is not supplied to the advance chamber 41 in the same manner as when the engine is stopped. The hydraulic pressure is discharged from 46 into the oil pan 48 to maintain a low pressure state.

  Therefore, the vane member 32 rotates counterclockwise in the drawing as shown in FIG. 8 against the spring force of the coil springs 55 and 56 as the pressure in the retard chamber 42 increases. As a result, the drive shaft 6 rotates relative to the timing sprocket 30 toward the retard side.

  For this reason, the closing timing of the intake valves 4 and 4 is delayed as shown by the solid line in FIG. 10, so that the valve overlap is reduced and combustion is improved. Thereby, idling rotation can be stabilized.

  Thereafter, when the vehicle starts traveling and shifts to a predetermined low-rotation load range, for example, the electromagnetic switching valve 47 is actuated by a control signal from the controller 39 to connect the supply passage 45 and the first hydraulic passage 43. On the other hand, the drain passage 46 and the second hydraulic passage 44 are communicated.

  Accordingly, the oil pressure in the retard chamber 42 is returned to the oil pan 48 from the drain passage 46 through the second hydraulic passage 44 this time, while the inside of the retard chamber 42 becomes low pressure, while in the advance chamber 41. The hydraulic pressure is supplied and the pressure becomes high.

  Therefore, the vane member 32 rotates clockwise in the drawing by the high pressure in the advance chamber 41 and the spring force of each coil spring 55, 56, and advances the relative rotational phase of the drive shaft 6 with respect to the timing sprocket 30. Convert to the side. On the other hand, the intake valves 4 and 4 are controlled to a slightly large operating angle by the variable lift mechanism 1. This increases the valve overlap between the intake valves 4 and 4 and the exhaust valves 5 and 5. For this reason, the pumping loss is reduced, and the fuel consumption can be improved.

Further, the rotation range usual in the low rotation region of the engine and with further shift to the high speed region, the vane member 32, as shown in FIG. 8, the hydraulic pressure supplied to the advance chamber 41 decreases, conversely The hydraulic pressure in the retard chamber 42 increases, and the relative rotational phase of the timing sprocket 30 and the drive shaft 6 is converted to the retard side against the spring force of the coil springs 55 and 56. Thus, in combination with the maximum lift control with the variable lift mechanism 1, the intake valves 4, 4 are sufficiently closed when the valve overlap between the intake valves 4, 4 and the exhaust valves 5, 5 is secured to some extent. Slowly, the intake efficiency (filling efficiency) of fresh air improves. This makes it possible to improve the engine output.

  Next, control for improving startability by the controller 22 in the present embodiment will be described with reference to FIG.

  First, in step 1, it is determined whether or not the ignition key switch is turned on. Here, if it is not turned on, the engine is in a stopped state, so the process returns as it is, but if it is turned on, the process proceeds to Step 2.

  In step 2, information signals such as the current engine speed N from the crank angle sensor 27, the engine body temperature T1 from the engine cooling water temperature sensor, and the atmospheric humidity H1 from the atmospheric humidity sensor are read, and the current engine state is read. Is detected.

  In step 3, the current operating position of the variable valve timing mechanism 2 is detected from the angular position information of the crank angle sensor 27, the drive shaft angle sensor 28, and the (absolute angle sensor).

  Subsequently, in step 4, the current operating position of the variable lift mechanism 1, that is, the operating angle (lift) is read from the control shaft angle sensor 29.

  In step 5, in order to obtain good startability, the target position of the lift control mechanism 1 is calculated by calculation based on the current operating position of the variable valve timing mechanism 2, and in step 6, the variable lift mechanism 1 is A current for switching to the position is output to the electric motor 20.

  That is, when the operating position of the variable valve timing mechanism 2 is deviated from the expected position due to malfunction or the like, the target position of the variable lift mechanism 1 is corrected in anticipation of this.

  Specifically, for example, as shown by the solid line a in FIG. 10, the opening / closing timing of the intake valves 4 and 4 is retarded from the target position (broken line a), and the closing timing (IVC) is the expected position (broken line). If it is on the retarding side from a), the effective compression ratio may be reduced and the startability may be hindered. Therefore, as shown by the solid line b in FIG. 10, the operating angle reduction control is performed by the variable lift mechanism 1, and the correction control is performed to the target position for advancing the IVC of the intake valves 4 and 4 to increase the effective compression ratio and the intake air. The amount is increased to improve the startability.

Further, since the mechanical friction increases when the temperature T1 of the engine body is low, the intake air amount (volume) required for starting is also increased. On the other hand, when the atmospheric humidity H1 is high, the torque decreases, so the intake air amount (volume) necessary for starting increases. A further correction value of the required intake air amount (volume) is determined based on the temperature T1 and the humidity H1, thereby further correcting the closing timing IVC. That is, the position of the variable lift mechanism 1 is calculated as a target value so that the ideal valve timing can be obtained by considering the startability on the premise of the current closing timing IVC position and further by the engine body temperature T1, the humidity H1, and the like. (Note that it is possible to consider parameters other than the closing timing IVC. Also, for example, it is possible to perform emission reduction control by valve overlap.)
Subsequently, in step 7, the current value I to the electric motor 20 is read, and in step 8, it is determined whether or not the current value I exceeds the peak.

  Here, if the current value I does not exceed the peak, the process returns to step 7. If it is determined that the current value I has exceeded, the process proceeds to step 9, where the starter motor 07 is energized to start cranking. That is, as shown in FIG. 11, energization of the starter motor 07 is started when the peak Tp of the current value I of the electric motor 20 of the variable lift mechanism 1 is exceeded. For this reason, since the peak Tp of the current value I of the electric motor 20 and the starter motor do not overlap, the load of the battery power source can be reduced.

  Next, in step 10, when cranking is started and the rotation of the crankshaft 02 is increased, complete explosion control such as fuel injection and ignition is performed, and smooth start is completed.

  FIG. 12 shows a control flowchart of the controller 22 in the second embodiment, and assumes a state in which a start-up vibration such as a warm-up start is likely to occur. From Step 11 to Step 14, as in Steps 1 to 4 shown in FIG. 9 of the first embodiment, when the ignition key switch is turned on, information signals from various sensors are read to detect the state of the engine. Then, the current position of the variable valve timing mechanism 2 and the position of the variable lift mechanism 1 are detected.

  In step 15, it is determined based on the current crank angle read from the crank angle sensor 27 whether or not the piston position of the cylinder in the intake compression stroke is more than a predetermined distance from the bottom dead center. Here, when it is determined that the distance is more than a predetermined difference, that is, when the crank stop position is the normal position Q as shown in FIG. 13, in step 16, the engine is operated by the method described in step 5 of the first embodiment. The target position of the variable lift mechanism 1 suitable for starting is calculated.

  If it is determined in step 15 that the position is not more than a predetermined distance and is located at or near the bottom dead center position (in the abnormal position Q1 in FIG. 13), the process proceeds to step 17.

  In step 17, the electric motor 20 is energized to control the target position of the variable lift mechanism 1 so that the operating angle of the intake valves 4 and 4 becomes a large operating angle Z as shown in FIG. That is, when the piston is located at or near the bottom dead center, the operating angle Z1 of the intake valves 4 and 4 is the small operating angle Z1, and the closing timing IVC of the intake valves 4 and 4 is decreased. Since the position is away from the dead center position, air enters the cylinder while the engine is stopped. If cranking is performed in this state, the compression may be excessive and large vibrations may occur during the initial cranking.

  Therefore, in the step 17, if the operating angle of the intake valves 4 and 4 is controlled to the target large operating angle Z by the variable lift mechanism 1 on the premise of the current position of the variable valve timing mechanism 2, the closing timing IVC is lowered. The target position is delayed from the dead center position. Therefore, the intake valves 4 and 4 are switched to the open state in the vicinity of the bottom dead center. Therefore, the closing after the corrected bottom dead center is performed at the time of the first cranking. Compression is started for the first time from time IVC. For this reason, the compression mentioned above falls and it can prevent a vibration.

  Next, in step 18, the electric motor 20 is energized for switching so that the target position of the variable lift mechanism 1 determined in step 16 or 17 is reached.

  Next, at step 19, the current operating position of the variable lift mechanism 1 switched by the operation of the electric motor 20 is detected by the control shaft angle sensor 29 and read.

  In step 20, it is determined whether or not the variable lift mechanism 1 has reached the target position. If it is determined that the target position has not been reached, the process returns to step 19, but if it is determined that it has reached, the process proceeds to step 21. To do.

  In step 21, the starter motor 07 is energized to start cranking. In step 22, fuel injection and ignition control are performed, and complete explosion control is performed at the valve timing suitable for the initial cranking. By this, it is possible to secure a higher accuracy starting performance.

  FIG. 14 shows a flowchart of the controller 22 in the third embodiment. This is not the position detection of the variable valve timing mechanism 2 performed before cranking, but is performed immediately before the previous engine stop. In this regard, the point that the drive shaft 6 is rotating is different from the other embodiments.

  That is, during the previous engine operation, the position of the variable valve timing mechanism 2 is appropriately detected by a trigger type sensor, and the data is stored in advance in the memory of the controller 22.

  Then, when the ignition key switch is turned off in step 21, in step 22, the current position of the variable valve timing mechanism 2 is immediately determined from the detection signal of the trigger type pickup sensor and the detection signal of the crank angle sensor 27. By comparison, the position of the variable valve timing mechanism 2 is detected and stored in the memory as described above. The position of the variable valve timing mechanism 2 detected by these is stored in the memory of the controller 22 even after the engine is stopped.

  In step 23, when the engine stops instantaneously as in the engine stall, the latest memory is stored as the position of the valve timing variable mechanism 2.

  Next, in step 24, it is determined whether or not the ignition key switch is turned on. If it is not turned on, the process directly returns. If it is turned on, the process proceeds to step 25.

  In step 25, the current engine state is detected by various sensors as described above. In step 26, the current position of the variable lift mechanism 1 is detected by the same method as described above. Then, in step 27, the previous engine is detected in step 22. The target position of the variable lift mechanism 1 is detected on the premise of the position of the variable valve timing mechanism 1 stored at the time of stop.

  Subsequently, in step 28, a current for switching control is output to the electric motor 20 of the variable lift mechanism 1 so as to operate to the target position. The following steps 29 to 32 including this step include the second step. The same processing as in steps 18 to 22 of the embodiment is performed.

  Therefore, in this embodiment, as described above, an apparatus using a trigger sensor that can be detected while having a time interval only during rotation without using an expensive absolute angle sensor as a position sensor of the valve timing variable mechanism 2. The manufacturing cost can be reduced.

  15 to 18 show a fourth embodiment, and a locking mechanism between the vane member 32 and the housing 34, which is a fixing means for restricting and releasing the rotation of the vane member 32 with respect to the housing 34. Is provided.

  That is, the locking mechanism is provided between the one vane 32b having a large width and the rear cover 36, and has a sliding hole 50 formed along the drive shaft 6 axial direction inside the vane 32b. A lid-shaped cylindrical lock pin 51 slidably provided in the sliding hole 50 and a cross-sectional cup-shaped engagement hole constituting portion 52 fixed in a fixing hole provided in the rear cover 36 are provided. The lock pin 51 is held by an engagement hole 52a in which the tapered tip 51a of the lock pin 51 engages and disengages, and a spring retainer 53 fixed to the bottom surface side of the sliding hole 50, and the lock pin 51 is engaged with the engagement hole 52a. And a spring member 54 biased in the direction.

  The engagement hole 52a is supplied with hydraulic pressure in the retard chamber 42 through an oil hole (not shown).

  When the engine is stopped, that is, when the engine is started, as shown in FIG. 16, the lock pin 51 is at a position where the vane member 32 is rotated to an approximately intermediate position between the most advanced angle and the most retarded angle, and the tip 51a is the The spring force of the spring member 54 is engaged with the engagement hole 52 a to lock the relative rotation between the timing sprocket 30 and the drive shaft 6. Therefore, the opening / closing phase (closing timing) of the intake valves 4 and 4 is held at an intermediate position closer to the advance side as shown by the broken line a in FIG. For this reason, an effective compression ratio increases and the startability at the time of a cold machine becomes favorable. In addition, emission is improved by moderate valve overlap.

  When the engine is started during the warm-up idling operation after the engine is started, the hydraulic pressure is supplied into the engagement hole 52a together with the hydraulic pressure supplied to the retard chamber 42, and the lock pin 51 resists the spring force of the spring 54. Then, the front end 51a is retracted from the engagement hole 52a and is disengaged, and the vane member 32 is relatively rotated to the most retarded angle side as shown in FIG. Therefore, the intake valves 4 and 4 are in the most retarded phase as shown by the solid line b in FIG. For this reason, valve overlap is reduced, and idling rotation can be stabilized.

  When the engine speed is shifted to the low engine speed range, as shown in FIG. 18, the hydraulic oil is discharged from the retard chamber 42 and supplied to the advance chamber 41. As a result of relative rotation to the corner side, the open / close phase of the intake valves 4 and 4 becomes the most advanced angle phase as shown by the broken line b in FIG. This reduces the pumping loss and improves the fuel efficiency.

  When the lock pin 51 cannot be engaged with the engagement hole 52a for some reason, for example, when the engine is stopped (during cold start), that is, when the lock pin 51 is in a position not suitable for start-up. In the same manner as described above, the controller 22 controls the operation of the variable lift mechanism 1 to correct the closing timing (IVC) of each of the intake valves 4 and 4 to the advance side or the retard side, and to the intermediate position suitable for starting. The corresponding IVC is controlled. As a result, the effective compression ratio is improved and the cold startability can be improved. In addition, emission can be improved by controlling the valve overlap.

  The present invention is not limited to the configuration of each of the embodiments described above. For example, the variable lift mechanism 1 is shown as the first variable mechanism, but any other mechanism can be used as long as it can be switched when the engine is stopped or cranked. For example, the variable valve timing mechanism described in Japanese Patent Application Laid-Open No. 2003-35115 may be any mechanism that can be switched when the engine is stopped by driving an electric motor.

  Moreover, although the said valve timing variable mechanism 2 was shown as a 2nd variable mechanism, the hydraulic drive type lift variable mechanism described in Unexamined-Japanese-Patent No. 2006-29245 may be sufficient, for example.

  Furthermore, in the present embodiment, an example in which the first variable mechanism and the second variable mechanism are provided side by side on the intake side is shown, but one variable mechanism may be provided on the intake side and the other variable mechanism may be provided on the exhaust side. Specifically, it is possible to provide a variable valve timing mechanism A driven by an electric motor on the intake side and a variable valve timing mechanism B driven hydraulically on the exhaust side. 9 is replaced with a variable valve timing mechanism A provided on the intake side, a variable valve timing mechanism 2 is replaced with a valve timing variable mechanism B provided on the exhaust side, and a variable valve timing mechanism before starting. Based on the position of B, for example, the variable valve timing mechanism A may be controlled before cranking starts or during cranking so that the target valve overlap amount is obtained.

  Further, in each of the above-described embodiments, the engine state temperature (cooling water temperature) and the crank angle have been described as the engine state. However, the temperature may be another part or another index. Further, atmospheric pressure other than atmospheric humidity may be atmospheric pressure that affects output torque.

  The technical ideas other than the invention described in the claims, as grasped from the embodiment, will be described below.

  (A) The variable valve operating apparatus for an internal combustion engine according to any one of claims 1 to 3, wherein the operating position of the second variable mechanism is detected immediately before cranking the engine.

  According to the present invention, since the operating position of the second variable mechanism is detected immediately before cranking, it is possible to accurately suppress the change in the operating position with time and to accurately detect the position, and to switch the first variable mechanism. It becomes possible to increase the accuracy of the target position.

  (B) The variable valve operating apparatus for an internal combustion engine according to any one of claims 1 to 3, wherein the operating position of the second variable mechanism is detected during the rotation before the previous engine stop.

  In this invention, since the operating position of the second variable mechanism is detected during the rotation before the previous engine stop, the detection sensor is not an expensive absolute angle sensor, but the second variable is determined from the timing when the predetermined phase passes with the rotation. An inexpensive trigger sensor for detecting the operating position of the mechanism can be provided. As a result, cost can be reduced.

  (C) a position for correcting the deviation of the target position of the first variable mechanism when the operating position of the second variable mechanism at the time of engine stop deviates from the target position set at the time of engine start. The variable valve operating apparatus for an internal combustion engine according to any one of claims 1 to 3, wherein the variable valve operating apparatus is set to be

  According to the present invention, when the operating position of the second variable mechanism is deviated from the target position (second target position) set at the time of starting the engine, the first target position of the first variable mechanism is determined as the deviation. Therefore, even when there is such a deviation, good startability can be ensured.

  (D) A variable valve operating system for an internal combustion engine according to any one of claims 1 to (c), wherein the state of the engine is detected by a crank angle before cranking.

  In this invention, since the state of the engine is detected by the actual crank angle before cranking, the deterioration of the engine starting performance caused by the difference in the crank angle that is likely to occur when the crank angle is not targeted for detection is suppressed. Can do.

  (E) When the crank angle in the cylinder during the intake stroke and the compression stroke is near the bottom dead center of the piston, the intake valve is opened at the target position of the first variable mechanism at the rotation angle of the crankshaft in the cylinder. The variable valve operating apparatus for an internal combustion engine according to (d), wherein the variable valve operating apparatus is set to a position.

  When the crank angle in the cylinder for the intake and compression strokes is near the bottom dead center of the piston, the first target position of the first variable mechanism is the position where the intake valve is open at this crank angle in the cylinder. During the first cranking, atmospheric pressure air accumulated in the cylinder can be discharged from the intake valve. Therefore, the reduction effect of noise and the effect of increasing the cranking rotation speed can be obtained by the decompression effect.

  (F) In addition to the operating position of the second variable mechanism and the state of the engine before cranking of the engine, the target operating position of the first variable mechanism is set according to the atmospheric state. The variable valve operating apparatus for an internal combustion engine according to any one of (c).

  According to the present invention, even if the atmospheric state at the time of engine start changes, good start performance can be obtained.

(G) Actuating the first variable mechanism electrically,
The cranking by the electric motor is started after the switching signal to the target position is output to the first variable mechanism prior to engine cranking and the current peak at the time of output is exceeded. The variable valve operating apparatus for an internal combustion engine according to any one of (f).

According to the present invention, since the cranking is started after the peak current of the operating current of the first variable mechanism is exceeded, the peak current generation time of the operating current of the first variable mechanism and the cranking of the electric motor to be cranked are determined. Since the peak current generation time of the operating current does not overlap, an increase in the burden on the battery can be suppressed.
(H) The internal combustion engine according to any one of claims 1, (a) to (g), wherein the first variable mechanism is provided on the intake side and the second variable mechanism is provided on the exhaust side. Variable valve gear.

  According to the present invention, for example, when the valve overlap is not suitable for starting, the first variable mechanism can correct it.

1 is a schematic diagram of an internal combustion engine provided for a first embodiment of a variable valve operating apparatus according to the present invention. It is a perspective view which shows the lift variable mechanism and valve timing variable mechanism which are provided to this embodiment. A and B are operation explanatory views at the time of small lift control by the variable lift mechanism. A and B are operation explanatory diagrams at the time of maximum lift control by the variable lift mechanism. It is a valve lift amount of an intake valve in this embodiment, an operating angle, and a valve timing characteristic view. It is sectional drawing of the valve timing variable mechanism with which this embodiment is provided. It is the sectional view on the AA line of FIG. 6 which shows the maximum advance angle control state by a valve timing variable mechanism. FIG. 7 is a cross-sectional view taken along line AA of FIG. 6 illustrating a maximum retard angle control state by a valve timing variable mechanism. It is a flowchart figure which shows the control by the controller of this embodiment. It is a characteristic view which shows the state which correct | amended the closing timing of the intake valve by the lift variable mechanism by a controller. It is a characteristic diagram of time and current value showing a state where the peak positions of energization of the electric motor and starter motor of the variable lift mechanism by the controller are shifted. It is a control flowchart figure by the controller of the 2nd Embodiment of this invention. It is a characteristic view showing a state where the closing timing of the intake valve is corrected by the operating angle control at the time of cranking by the variable lift mechanism of the present embodiment. It is a control flowchart figure by the controller of 3rd Embodiment. It is a longitudinal cross-sectional view of the valve timing variable mechanism which shows 4th Embodiment. It is the BB sectional view taken on the line of FIG. 15 which shows the intermediate position control state in this embodiment. FIG. 16 is a cross-sectional view taken along the line BB in FIG. 15 illustrating the most retarded angle control state in the present embodiment. FIG. 16 is a cross-sectional view taken along the line BB in FIG. 15 illustrating the most advanced angle control state in the present embodiment. It is a characteristic view which shows the opening / closing phase of the intake valve in each said control state.

Explanation of symbols

01 ... Piston 02 ... Crankshaft 07 ... Starter motor 1 ... Variable lift mechanism (first variable mechanism)
2. Valve timing variable mechanism (second variable mechanism)
4 ... Intake valve 6 ... Drive shaft 20 ... Electric motor 22 ... Controller 27 ... Crank angle sensor 28 ... Drive shaft angle sensor 29 ... Control shaft angle sensor

Claims (1)

  1. A first variable mechanism capable of controlling the operating characteristics of the engine valve during engine stop or cranking;
    A second variable mechanism that controls the operating state of the engine valve separately from the first variable mechanism,
    Depending on the operating state of the cranking start before the second variable mechanism of the engine, cranking before or during cranking, the operation of the first variable mechanism so as to satisfy a predetermined start condition of the engine A variable valve operating apparatus for an internal combustion engine, which outputs a control signal to be controlled prior to cranking and starts cranking with a predetermined delay time from a peak value of a control current of the first variable mechanism .
JP2006161760A 2006-06-12 2006-06-12 Variable valve operating device for internal combustion engine Expired - Fee Related JP4776447B2 (en)

Priority Applications (1)

Application Number Priority Date Filing Date Title
JP2006161760A JP4776447B2 (en) 2006-06-12 2006-06-12 Variable valve operating device for internal combustion engine

Applications Claiming Priority (4)

Application Number Priority Date Filing Date Title
JP2006161760A JP4776447B2 (en) 2006-06-12 2006-06-12 Variable valve operating device for internal combustion engine
US11/808,320 US7802546B2 (en) 2006-06-12 2007-06-08 Variable valve actuating apparatus and process for internal combustion engine
DE200710027076 DE102007027076A1 (en) 2006-06-12 2007-06-12 Variable valve actuator for an internal combustion engine
CN 200710110037 CN101089372A (en) 2006-06-12 2007-06-12 Variable valve actuating apparatus and process for internal combustion engine

Publications (2)

Publication Number Publication Date
JP2007332780A JP2007332780A (en) 2007-12-27
JP4776447B2 true JP4776447B2 (en) 2011-09-21

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US7802546B2 (en) 2010-09-28

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